This invention relates to gearbox assemblies for electric power steering assemblies.
Electric power steering systems use an electric motor to produce an assistance torque that is applied to a rotating part of the steering system. In a conventional arrangement this torque assists the driver in turning the wheel. Because motors work best at relatively high speeds and because compact motors produce relatively low torques, the connection between the output of the motor and the steering column is usually through a reduction gearbox.
The most widely used type of electric power assisted steering reduction gearboxes are of a relatively simple worm and gear configuration similar to that shown in FIG. 1 of the accompanying drawings The gearbox typically comprises a gearbox housing which houses a worm shaft and a gear wheel. The worm shaft is connected to the output of an electric motor. The motor may be secured to an end face of the housing or even located within the housing. The worm shaft is supported by a main bearing at an end closest to the motor and a tail bearing at an end furthest from the motor, both bearings typically comprising ball bearings supported within an inner bearing race that is threaded onto the worm shaft and an outer bearing race that is secured to the housing. The function of the bearings is to allow the worm shaft to rotate whilst to a certain degree limiting axial and radial movement as will be explained. The gear wheel is connected to an output shaft of the gearbox and located so that teeth of the gear wheel engage teeth of the worm shaft.
By choosing appropriate design parameters, this type of gearbox can be made to provide a large speed reduction ratio within compact dimensions. They use a single gear set having a low sliding friction coefficient between the teeth, typically less than 0.05, so that they are “back-driveable”. I.e., so that the gear can drive the worm almost as efficiently as the worm can drive the gear. This last attribute is essential for the safety of EPS systems of the kind in which the steering wheel is physically connected to the road wheels enabling some steering effect in the event of a fault at the motor or any other part of the motor control and drive system.
It is known that the speed reduction gearboxes used in electrical power-assisted steering (EPS) apparatus are prone to rattle due to external torsional vibrations acting at their output shafts. These vibrations originate at the road wheels due to surface roughness or unbalanced wheels. Alternatively, mechanical noise can arise from sudden torque reversals applied at the steering wheel by the driver.
The main rattle sites in a worm and wheel gearbox are shown in FIG. 1:—                (a) at the engagement of the worm and gear teeth and        (b) at the “main” ball bearing which axially locates the worm shaft.        
A well-known solution to the rattle at site (a), namely the so-called “Sprung Worm” mechanism, tends to exacerbate the rattle problem at site (b). An example of such a prior art arrangement is shown in FIG. 1 of the accompanying drawings. In the “Sprung Worm” mechanism, a biasing means such as a leaf spring applies a biasing force that urges the worm shaft into engagement with the wheel gear, the biasing force being chosen such that at low gearwheel torques a dual-flank contact of the worm and gearwheel teeth is achieved while ensuring that there is a transition to single-flank contact between the worm and wheel gear at higher gearwheel torques.
The biasing means requires a small amount of radial movement of the worm shaft and this is achieved by allowing it to pivot around its axis in the plane of the gearwheel by a small angle (typically less than +/−0.5 degrees) either side of its nominal position around an axis which is nominally at the centre of the main bearing at (b). This movement is controlled by                (i) a specially configured tail bearing at site (c) that is allowed to move by small amounts (typically less than +/−0.5 mm) in a vertical guiding device such as a linear bearing, and        (ii) the main bearing at (b) has sufficient internal axial clearance between its balls and the sides of its race grooves to permit a small articulation (i.e. tilting) angle which is typically less than +/−0.5 degrees.        
Unfortunately, the said axial clearance in the bearing at (b) inevitably allows relative axial free play between the inner and outer races of up to +/−0.2 mm, promoting rattle noises from the bearing when there are reversing torque impulses at the [[O]] output [[S]] shaft. These rattle noises can be partly, but not entirely, suppressed by applying a heavy axial biasing force to the worm shaft.
As stated earlier, if dual-flank contact can be maintained for gearwheel torque levels of up to a certain level, say 4 Nm, then gear rattle will not be serious problem in general driving. The said 4 Nm is referred to as the “kick-out” torque because it is the gearwheel torque at which the “normal” component of the tooth flank contact force is just sufficient to overcome the inward action of the biasing means and thereby causes the worm shaft to pivot outwards. The outward pivoting movement of the worm shaft is limited by the travel allowed at the tail bearing by its vertical guide. The downward biasing force required on the worm shaft tail bearing to achieve a 4 Nm kick-out torque is typically around 12N, equating to around 20N of downward force at the middle of the worm. The kick-out torque is the same for both directions of torque transmission only if the main bearing is pivoted in the tooth contact plane. If the worm shaft is pivoted at the main bearing centre, as shown in FIG. 1, or in any horizontal plane parallel to the nominal worm axis other than that of the tooth contact point, the kick-out torque with be different for the two directions of torque transmission due to the moments created by the axial component of the force resulting from the gearwheel torque.
A significant disadvantage of employing the above described “sprung-worm” method of gear rattle suppression is that the biasing means acting on the tail bearing causes an extra amount quiescent (i.e. background) friction in the gearbox. The increase is typically around a 0.5 Nm compared with a so-called “fixed-centre gearbox” having no worm shaft articulation. This increase is due to the extra rubbing friction which is created between the flanks of the worm and gearwheel teeth when they are forced into mesh by the ARS. The effect is exacerbated both by the “wedge” shape of the teeth, which is a function of their slope or “pressure angle”, and by the fact that extra rotational drag felt at the worm shaft is multiplied by the gearbox ratio (typically around 20:1) when measured at the gearwheel axis.